1. Field of the Invention
The invention is directed to a reduction of NOx and PM emissions from combustion engines. The field of application is primarily in internal combustion engines for motor vehicles, but this invention can also be utilized in other energy conversion xe2x80x9cenginesxe2x80x9d which utilize combustion of chemical fuels, including electric power generation plants.
2. The Prior Art
The growing use of motor vehicles greatly adds to the atmospheric presence of pollutants, such as oxides of nitrogen and particulate matter, and has created a demand for a significant reduction in such emissions.
Prior art gasoline engines generally operate with charge-air throttling and intake port fuel injection to provide a mixture of fuel and charge-air for inducting into the combustion chambers. The term xe2x80x9ccharge-airxe2x80x9d as used herein means either air or a mixture of air and recirculated exhaust gas. Charge-air throttling is used to control the load (or torque) output of the engine and results in significant efficiency penalties, especially at lower loads. Port fuel injection is used to provide good control and mixing of the required fuel with charge-air. Pre-mixed fuel and charge-air will auto-ignite during compression, depending on the fuel and charge-air characteristics, at a certain compression ratio which corresponds to the auto-ignition temperature of the fuel and charge-air mixture. Prior art gasoline engines are generally limited to compression ratios of between 9:1 and 10:1 to avoid uncontrolled auto-ignition. The ignition process is initiated by the sparking of a spark-plug so that rapid combustion begins at or near piston top dead center TDC, (generally between TDC and 20 crank angle degrees after TDC), and combustion propagation proceeds from the ignition location as a xe2x80x9cflame-frontxe2x80x9d traveling through the combustible mixture. At higher compression ratios and some operating conditions, the fuel and charge-air mixture auto-ignites in an uncontrolled manner and exhibits unacceptable xe2x80x9cknock.xe2x80x9d Not being able to safely operate at higher compression ratios significantly reduces the engine""s efficiency potential.
some gasoline engines utilize direct fuel injection (fuel is injectec directly into the combustion chanber), with or without charge-air throttling. Generally, these engines operateat low loads throgh stratified combustion. The fuel is injected relatively late in the compression stroke with little or no charge-air throttling. A spark then initiates combustion that can occur as long as the stratified mixture is within the flamability limits of the fuel. Since the late injection allows less time for good fuel and charge-air mixing than pre-mix operation, such combustion is characterized by higher unburned fuel and particulate emissions. Also, localized temperatures are high and NO is formed and becomes part of the exhaust emissions. At higher loads the beginning of the fuel injection occurs earlier to allow more time for fuel and charge-air mixing. The earlier fuel injection limits the compression ratio for gasoline to levels comparable to the pre-mixed engines since in effect they become pre-mixed engines at high loads. Pre-mixed gasoline engines also experience high combustion temperatures and produce significant NO emissions.
Prior art diesel engines operate over all loads with late direct fuel injection and with little or no charge-air throttling. Diesel engines also operate at a relatively high compression ratio (generally between 15 and 20 to one) as compared to prior art gasoline engines because they make use of the auto-ignition properties of diesel fuel. Diesel fuel will under all intended operating conditions auto-ignite when injected into the compressed charge-air at or near TDC. As a result of these operating characteristics, diesel engines exhibit high efficiency. The primary problems with diesel engines are the unburned fuel, particulate and NOx emissions, as previously explained for late direct injection of gasoline. Gasoline could be used in prior art diesel engines by adding an assured source of ignition (e.g., a spark or glow plug), but still with the same emission problems.
Accordingly, it is an object of the present invention to provide for both efficient engine operation and extremely low levels of NOx emissions.
It is another object of the present invention to provide a method of engine operation wherein compression temperature and peak combustion temperature are controlled over an engine""s speed and load range.
Accordingly, the present invention provides a method of operation of an internal combustion engine including intaking ambient air through at least one compressor to provide a charge-air at a boosted pressure and introducing the boosted charge-air to the internal combustion engine. The method further includes introducing fuel into the internal combustion engine for combustion in admixture with the charge-air at a combustion temperature approximating a target value, producing an exhaust gas. Engine operating conditions, inclusive of torque demand, are sensed and the boosted pressure is changed proportional to a change in the sensed torque demand so as to maintain the combustion temperature at approximately the target value, which value is below 2100xc2x0 K.
In a preferred embodiment the method further involves passing a portion of the charge-air through a heat exchanger with bypass of the heat exchanger by a second portion of the charge-air. The temperature of the charge-air is sensed downstream of the bypass line and heat exchanger and the amount of the second portion bypassing the heat exchanger is controlled by operation of a control valve in the bypass line to bring the sensed intake temperature to a target temperature determined in accordance with the sensed engine operating conditions.
A portion of the exhaust gas may be recirculated for admixture with the charge-air and fuel. In this case, oxygen concentration in the admixture is sensed and the amount of EGR is regulated to bring the sensed oxygen concentration to a target oxygen concentration determined for the sensed engine operating conditions. The fuel feed is controlled responsive to the sensed temperature of the charge-air intake and the sensed boost pressure.
The fuel may be introduced into the charge-air downstream or upstream of the compressor so that the fuel is contained within the charge-air.
Thus, the present invention provides, in an internal combustion engine, high compression ratio (e.g., generally greater than 15 to 1 in the compression stroke) operation with little or no charge-air throttling, as is characteristic of the high-efficiency diesel-cycle engine, but without the emission problems of prior art engines, through a unique, new method of operation.
The formation rate of the pollutant NO during the fuel combustion process can be generally expressed in simplified form as follows:
NO formation rate=C1[N2]C2[O2]C3 expC4Txe2x80x83xe2x80x83(I)
Where: C1, C2, C3, and C4 (Cx) are constants, [N2] is the concentration of nitrogen, [O2] is the concentration of oxygen, exp is a constant, and T is the absolute temperature of the mixture.
Since temperature is an exponent in the above formula (I), it can be expected that for given concentrations of nitrogen and oxygen, the NO formation rate increases exponentially with temperature. This relationship is widely recognized and is shown graphically in FIG. 1 for typical engine operating conditions. Engine combustion times general fall within one to five milliseconds. It can clearly be seen that the formation of NO will be minimal if the engine combustion temperature can be maintained below about 2000 degrees Kelvin(K). The desirability of maintaining combustion temperatures below this level and yet still have combustion rapid enough to be complete for practical engine speeds, is well known. A recent Society of Automotive Engineers Technical Paper (#2000-01-1177) by Patrick F. Flynn and others from Cummins Engine Company reflects prior art understanding of the known methods of operation of engines with the goal of controlling combustion temperature to reduce formation of NO. This paper concludes that, for gasoline and other spark-ignition engines, xe2x80x9cthe minimum possible peak combustion temperature is 2100 Kxe2x80x9d and xe2x80x9cNOx numbers show a limit of 0.5 g/bhp-hr,xe2x80x9d and that, for diesel engines, xe2x80x9cthe lowest possible peak combustion temperature would be approximately 2300 Kxe2x80x9d with a xe2x80x9cNOx emissions level of 1.0 g/bhp-hr.xe2x80x9d
The present invention provides a new method of engine operation that yields stable and efficient combustion at temperatures below 2100 K. Results for NOx emissions with gasoline, diesel and other fuels are consistently less than 0.2 g/bhp-hr, substantially below that for prior art engines.
Referring again to the Equation I for NO formation rate, the oxygen concentration must be sufficient to fully react with the available fuel for a given engine operating condition, e.g., charge-air boost level. The nitrogen concentration is naturally high in the charge-air and, therefore, temperature is the one variable available for control to limit NO formation, once the oxygen concentration is minimized and optimized for the operating condition. It is also critical to control local temperature as NO is rapidly formed wherever the temperature is above 2000 K.
In control of the peak temperature of combustion by the method of the present invention two factors are most important. First, the temperature, T1, of the charge-air, or the charge-air fuel mixture if fuel is pre-mixed, at the beginning of compression, must be controlled. Generally, the objective is to minimize T1, since the compression process is a multiplier of T1 (in degress absolute). For an ideal gas, the final compression temperature, T2, is a function of the compression ratio, CR, assuming adiabatic compression, i.e., T2=T1 f(CR) (where f(CR) is a function of CR). For example, for a compression ratio of 16, the multiplier of T1 is about 3. Therefore, if the T1, were 300xc2x0 Kelvin (27xc2x0 Celsius), T2 would be 900 K. However, if T1 were 400xc2x0 Kelvin (127xc2x0 Celsius), T2 would be 1200 K.
Second, for a given quantity of fuel to be burned, assuming adiabatic combustion, the final combustion temperature, T3 can be calculated as follows: T3=T2+Hc/Cv (where Hc=heat released from combustion of the fuel and Cv is the total heat capacity of the charge-air fuel mixture, i.e., the mass of the mixture times the specific heat capacity). Since for a given quantity of fuel burned, Hc is fixed, the only variable available for control of T3 is Cv. If Cv is large, T3 will be lower. The quantity of fuel burned, in turn, is a function of (proportional to) torque demand.
The present invention controls T3, for example to 2000xc2x0 Kelvin (see FIG. 1), to minimize NO formation. Therefore, for a constant T2, Hc/Cv must be maintained constant. To maintain Hc/Cv constant, Cv must increase as the quantity of fuel combusted (engine load) is increased. Since Cv is of the form Cv=cvM (where cv is the specific heat capacity of the charge-air fuel mixture and M is the mass of the charge-air fuel mixture), M must be increased as the quantity of fuel combusted is increased and decreased as the quantity of fuel combusted is decreased. This is accomplished in the present invention by controlling the boost pressure of the charge-air in the intake system, i.e., controlling the charge-air density. The mass M is proportional to the pressure of the charge-air.
Another important factor to consider and account for is the fact that a real engine is not adiabatic. If the temperature of the charge-air as it enters the intake system is To, and is at a temperature lower than portions of the intake system, heat will flow from the intake system into the charge-air and its temperature will rise. Also, as the charge-air is inducted past the intake valve(s) into the engine cylinder, it will be exposed to the hot surfaces of the cylinder head, piston top and cylinder walls, surfaces that were just prior exposed to the combustion process and the hot (e.g., 2000xc2x0 at the end of combustion) combustion gases during the period of expansion and exhaust strokes (for a four-stroke engine). Therefore, heat will flow into the charge-air during the intake stroke and early portions of the compression stroke. To control (usually minimize) the rise in temperature of the charge-air prior to the compression temperature rise, the method of operation of the present invention primarily utilizes control of the boost pressure. As the boost pressure controls the mass of charge-air (or if fuel is present, the charge-air/fuel mixture, both also referred to herein as xe2x80x9ccharge massxe2x80x9d or simply xe2x80x9cchargexe2x80x9d), it directly controls the temperature rise of the charge mass since, for a given heat energy flow into the charge mass from the system surfaces, the temperature rise is directly proportional to the mass of the charge, as shown by the relationship: T1=T0+Hw/Cv, where Hw is the heat energy from the system surfaces. Cv is increased to reduce T1 by increasing charge mass as previously shown.
The method of operation of the first and second embodiments of the present invention controls T2 to assure that auto-ignition and combustion are sufficiently rapid to be efficient and effective over the operating speed and load of practical engines (generally combustion that is 90% complete within 1 to 5 milliseconds), and controls T3 to assure minimal NO formation. Controlling T3 to the levels of the present invention also improves engine efficiency since the lower than prior art T3 levels reduce the heat loss from the combustion gases during expansion. Heat (energy) that flows through the system walls to the engine xe2x80x9ccoolantxe2x80x9d would, if retained in the combustion gases, be used to sustain higher system pressures during expansion and thus extract more useful work for a given quantity of fuel burned. With a lower T3 the temperature difference, xcex94T, between the combustion gases and the system surfaces is lower and less heat energy flows to the coolant.